This invention relates generally to multistage turbocompressors of the type having a plurality of impellers or fan wheels mounted on a single rotating shaft and operating to compress a gaseous fluid such as air or a gas. More specifically, the invention relates to a multistage turbocompressor of the above stated character in which all or most of the impellers are of the diagonal-flow or "mixed-flow" type with exit flow angles increased from the impellers near the suction end toward those near the discharge end thereby to cause the specific (rotational) speed of each impeller to be within its optimal range.
In general, a gaseous fluid such as air or a gas possesses compressibility, and, therefore, when the gaseous fluid is compressed for the purpose of raising its pressure, its volume decreases according to Boyle's law (also known as Mariotte's law) as is well known. For a 4-stage compressor to suck in air and produce a discharge or delivery pressure of 7 kg.f/cm.sup.2.G, it is necessary that the pressure ratio (i.e., the ratio of the absolute discharge and suction pressures) of each stage be selected at a value of the order of 1.7, and the volumetric flow rate of the gaseous fluid sucked into a fan wheel or impeller is reduced approximately 60 percent upon reaching the entrance of the succeeding stage.
For the purpose of obtaining a discharge or delivery pressure of 7 kg.f/cm.sup.2.G with 3-stage compression, it is necessary to select a pressure ratio of approximately 2 for each stage. In this case, the volumetric flow rate at the entrance of the impeller of the succeeding stage is decreased to approximately 50 percent of that of the preceding stage. Thus, as the pressure ratio per stage increases, the rate of decrease of the volumetric flow rate of the gaseous fluid sucked into the impeller of a succeeding stage increases.
On one hand, in order for the impeller of each stage to exhibit high efficiency, it is required that the specific (rotational) speed N.sub.s expressed by the following equation be within an optimum range for each stage. EQU N.sub.s =N.multidot.Q.sup.1/2 /H.sub.ad .sup.3/4, (1)
where: N is the impeller rotational speed (r.p.m.); Q is the volumetric flow rate (m.sup.3 /min.) of each stage; and H.sub.ad is the adiabatic head (m.) of each stage. This specific speed N.sub.s is derived from the fluid mechanical law of similarity of turboblowers and compressors. It is a quantity having an important relation to the performance of the turbomachine and is an essential factor also in the selection of the type of the impellers.
Among the types of impellers, the common types are the centrifugal type, the diagonal-flow or "mixed-flow" type, and the axial-flow or propeller type. For each type, there is an optimum specific speed, and impellers of equal specific speed N.sub.s are geometrically similar impellers irrespective of their sizes and their rotational speeds. Furthermore, the optimum value of the specific speed N.sub.s has the characteristic of increasing with increasing width of the impeller blades in the centrifugal type and, further, with transformation into the diagonal-flow type.
Heretofore, in multistage turbocompressors, the impellers of the multiple stages have been of the axial-flow type, the centrifugal type, or a combination of the two types. For example, in one common type, a plurality of centrifugal type impellers are successively fixed in tandem on a single rotating shaft, and all impellers have the same outer diameter. When the outer diameters of the impellers are thus made all equal, the specific speed N.sub.s decreases in proportion to .sqroot.Q toward the rear stages. Accordingly, the shapes of the impellers are so designed that, on the low-pressure side including the initial stage, the ratio of the inner and outer diameters of each impeller is large, and the blade width is wide, while, on the high-pressure side, the ratio of the inner and outer diameters becomes smaller and the blade width becomes narrower while the blade flow path becomes longer as the final stage is approached.
However, the range of optimal specific speed wherein a centrifugal type impeller can exhibit high efficiency is not very wide. For this reason, in general, the specific speed N.sub.s of only the impellers of one part become optimal, and decrease in the efficiencies of the other impellers cannot be avoided. This has been the cause of lowering of the efficiency of the compressor as a whole.
In one example of a known multistage turbocompressor which has been designed to overcome this problem as much as possible, a plurality of impellers are successively fixed in tandem on a single rotating shaft, as in the above described example, but the initial-stage impeller is a doulbe-suction impeller which sucks gaseous fluid at opposite axial ends. It is contemplated by this arrangement to increase the capacity of the initial-stage impeller, to make the range of the specific speeds N.sub.s of the other impellers as narrow as possible, and to increase the efficiency of the compressor as a whole. In this case, however, the flow path from the double-suction impeller of the initial stage to the entrance of the impeller of the succeeding stage becomes complicated and gives rise to problems in the construction of the compressor casing.
In still another example of a known multistage turbocompressor, the achievement of the same object as in the preceding example is comtemplated by providing an initial-stage impeller having a wide, unobstructed entrance at the front end of the rotating shaft, which, at its front end part is a cantilever shaft. Thus, the capacity of the initial-stage impeller is made large. The optimal specific speed N.sub.s thereof is made high, and the range of the specific speeds N.sub.s of the impellers of the succeeding stages is made narrow. In this case, however, since a front or inboard bearing is required between the impeller of the initial stage and that of the succeeding second stage, the compressor casing is divided at the part between the initial stage and the remainder of the stages, which gives rise to complications in the construction. Furthermore, the diameter of the rotating shaft at the front bearing must be made greater for the sake of strength, whereby there arise problems such a severe design conditions due to the resulting increase in the shaft circumferential velocity.
The overcoming of these problems was evidently contemplated in a further example of a known multistage turbocompressor, in which a plurality of impellers, which are fixed successively in tandem on one rotating shaft, are divided into three impeller groups seuqentially from the front or suction end to the rear or discharge end. The outer diameters of the impellers in each group are the same, but the outer diameters of the impellers of the three groups are successively decreased from the front end to the rear end, and the specific speed N.sub.s of each impeller is made to approach as close to the optimum value as possible. In this case, while the efficiencies of all impellers can be made high, since the outer diameters of the impellers of the rearward groups are small, their circumferential velocities are low, and the pressure head per stage decreases as the second power of the circumferential velocity.
That is, the pressure ratios of the impellers of the rearward stages decrease, and the compressing capacity of the compressor as a whole decreases. Alternatively, in order to obtain the same compressing capacity, it is necessary to increase the number of stages beyond that ordinarily used in the above described known compressore. Furthermore, since the centrifugal stress produced in an impeller also decreases as the square of th circumferential velocity, the centrifugal stress imparted to the impellers of the rear stages becomes much lower than the allowable stress of the impeller material, whereby the efficiency of material utilization drops remarkably.
As described above, in the case where all of the impellers of a known multistage turbocompressor are of the centrifugal type, the efficiency of the compressor as a whole decreases, and attempts to prevent this decrease in the efficiency have unavoidably entailed problems such as complication of construction and necessity of increasing the number of stages.